Friction element load sensing in an automatic transmission

ABSTRACT

A load sensor assembly for measuring an amount of torque transmitted through a torque establishing element includes a core mounted on a transmission housing and a load sensor mounted on the core. The load sensor is positioned against a portion of the torque establishing element whereby a portion of the amount of torque transmitted through the torque establishing element travels through the load sensor and is measured. A cable is connected to the load sensor for transmitting a signal representative of the amount of torque to a transmission controller.

FIELD OF INVENTION

The present invention pertains to the field of automatic transmissionsfor motor vehicles and, more particularly, to a friction element loadsensor that directly measures torque transmitted by a friction elementof an automatic transmission.

BACKGROUND OF THE INVENTION

A step-ratio automatic transmission system in a vehicle utilizesmultiple friction elements for automatic gear ratio shifting. Broadlyspeaking, these friction elements may be described as torqueestablishing elements although more commonly they are referred to asclutches or brakes. The friction elements function to establish powerflow paths from an internal combustion engine to vehicle tractionwheels. During acceleration of the vehicle, the overall speed ratio,which is the ratio of a transmission input shaft speed to a transmissionoutput shaft speed, is reduced during a ratio upshift as vehicle speedincreases for a given engine throttle setting. A downshift to achieve ahigher speed ratio occurs as an engine throttle setting increases forany given vehicle speed, or when the vehicle speed decreases as theengine throttle setting is decreased.

Various planetary gear configurations are found in modern automatictransmissions. However the basic principle of shift kinematics remainssimilar. Shifting a step-ratio automatic transmission having multipleplanetary gearsets is accompanied by applying and/or releasing frictionelements to change speed and torque relationships by altering the torquepath through the planetary gearsets. Friction elements are usuallyactuated either hydraulically or mechanically.

In the case of a synchronous friction element-to-friction elementupshift, a first pressure actuated torque establishing element, referredto as an off-going friction element, is released while a second pressureactuated torque establishing element, referred to as an on-comingfriction element, engages in order to lower a transmission gear ratio. Atypical upshift event is divided into preparatory, torque and inertiaphases. During the preparatory phase, an on-coming friction elementpiston is stroked to prepare for its engagement while an off-goingfriction element torque-holding capacity is reduced as a step toward itsrelease. During the torque phase, which may be referred to as a torquetransfer phase, on-coming friction element torque is raised while theoff-going friction element is still engaged. The output shaft torque ofthe automatic transmission typically drops during the torque phase,creating a so-called torque hole. When the on-coming friction elementdevelops enough torque, the off-going friction element is released,marking the end of the torque phase and the beginning of the inertiaphase. During the inertia phase, the on-coming friction element torqueis adjusted to reduce its slip speed toward zero. When the on-comingfriction element slip speed reaches zero, the shift event is completed.

In a synchronous shift, the timing of the off-going friction elementrelease must be synchronized with the on-coming friction element torquelevel to deliver a consistent shift feel. A premature release leads toengine speed flare and a deeper torque hole, causing perceptible shiftshock for a vehicle occupant. A delayed release causes a tie-up of gearelements, also resulting in a deep and wide torque hole for inconsistentshift feel. A conventional shift control relies on speed measurements ofthe powertrain components, such as an engine and a transmission inputshaft, to control the off-going friction element release process duringthe torque phase. A conventional torque phase control method releasesthe off-going friction element from its locked state through anopen-loop control based on a pre-calibrated timing, following apre-determined off-going friction element actuator force profile. Thisconventional method does not ensure optimal off-going friction elementrelease timing and therefore results in inconsistent shift feel.

Alternatively, a controller may utilize speed signals to gauge off-goingfriction element release timing. That is, the off-going friction elementis released if the controller detects a sign of gear tie-up, which maybe manifested as a measurable drop in input shaft speed. When a releaseof the off-going friction element is initiated prematurely before theon-coming friction element develops enough torque, engine speed orautomatic transmission input shaft speed may rises rapidly in anuncontrolled manner. If this so-called engine speed flair is detected,the controller may increase off-going friction element control force toquickly bring down automatic transmission input speed or off-goingfriction element slip speed. This speed-based or slip-based approachoften results in a hunting behavior between gear tie-up and engineflair, leading to inconsistent shift feel. Furthermore, off-goingfriction element slip control is extremely difficult because of its highsensitivity to slip conditions and a discontinuity between static anddynamic frictional forces. A failure to achieve a seamless slip controlduring the torque phase leads to undesirable shift shock.

In the case of a non-synchronous automatic transmission, the upshiftingevent involves engagement control of only an on-coming friction element,while a companion clutching component, typically a one-way coupling,automatically disengages to reduce the speed ratio. The non-synchronousupshift event can also be divided into three phases, which may also bereferred to as a preparatory phase, a torque phase and an inertia phase.The preparatory phase for the non-synchronous upshift is a time periodprior to the torque phase. The torque phase for the non-synchronousshift is a time period when the on-coming friction element torque ispurposely raised for its engagement until the one-way coupling startsslipping or overrunning. This definition differs from that for thesynchronous shift because the non-synchronous shift does not involveactive control of a one-way coupling or the off-going friction element.The inertia phase for the non-synchronous upshift is a time period whenthe one-way coupling starts to slip, following the torque phase.According to a conventional upshift control, during the torque phase ofthe upshifting event for a non-synchronous automatic transmission, thetorque transmitted through the oncoming friction element increases as itbegins to engage. A kinematic structure of a non-synchronous upshiftautomatic transmission is designed in such a way that torque transmittedthrough the one-way coupling automatically decreases in response toincreasing oncoming friction element torque. As a result of thisinteraction, the automatic transmission output shaft torque drops duringthe torque phase, which again creates a so-called “torque hole.” Beforethe one-way coupling disengages, as in the case previously described, alarge torque hole can be perceived by a vehicle occupant as anunpleasant shift shock. An example of a prior art shift controlarrangement can be found in U.S. Pat. No. 7,351,183 hereby incorporatedby reference.

A transmission schematically illustrated at 2 in FIG. 1 is an example ofa prior art multiple-ratio transmission with a controller 4 whereinratio changes are controlled by friction elements acting on individualgear elements. Engine torque from vehicle engine 5 is distributed totorque input element 10 of hydrokinetic torque converter 12. An impeller14 of torque converter 12 develops turbine torque on a turbine 16 in aknown fashion. Turbine torque is distributed to a turbine shaft, whichis also transmission input shaft 18. Transmission 2 of FIG. 1 includes asimple planetary gearset 20 and a compound planetary gearset 21. Gearset20 has a permanently fixed sun gear S1, a ring gear R1 and planetarypinions P1 rotatably supported on a carrier 22. Transmission input shaft18 is drivably connected to ring gear R1. Compound planetary gearset 21,sometimes referred to as a Rayagineaux gearset, has a small pitchdiameter sun gear S3, a torque output ring gear R3, a large pitchdiameter sun gear S2 and compound planetary pinions. The compoundplanetary pinions include long pinions P2/3, which drivably engage shortplanetary pinions P3 and torque output ring gear R3. Long planetarypinions P2/3 also drivably engage short planetary pinions P3. Shortplanetary pinions P3 further engage sun gear S3. Planetary pinions P2/3,P3 of gearset 21 are rotatably supported on compound carrier 23. Ringgear R3 is drivably connected to a torque output shaft 24, which isdrivably connected to vehicle traction wheels through a differential andaxle assembly (not shown). Gearset 20 is an underdrive ratio gearsetarranged in series with respect to compound gearset 21. Typically,transmission 2 preferably includes a lockup or torque converter bypassclutch, as shown at 25, to directly connect transmission input shaft 18to engine 5 after a torque converter torque multiplication mode iscompleted and a hydrokinetic coupling mode begins. FIG. 2 is a chartshowing a clutch and brake friction element engagement and releasepattern for establishing each of six forward driving ratios and a singlereverse ratio for transmission 2.

During operation in the first four forward driving ratios, carrier P1 isdrivably connected to sun gear S3 through shaft 26 and forward frictionelement A. During operation in the third ratio, fifth ratio and reverse,direct friction element B drivably connects carrier 22 to shaft 27,which is connected to large pitch diameter sun gear S2. During operationin the fourth, fifth and sixth forward driving ratios, overdrivefriction element E connects turbine shaft 18 to compound carrier 23through shaft 28. Friction element C acts as a reaction brake for sungear S2 during operation in second and sixth forward driving ratios.During operation of the third forward driving ratio, direct frictionelement B is applied together with forward friction element A. Theelements of gearset 21 then are locked together to effect a directdriving connection; between shaft 28 and output shaft 26. The torqueoutput side of forward friction element A is connected through torquetransfer element 29 to the torque input side of direct friction elementB, during forward drive. The torque output side of direct frictionelement B, during forward drive, is connected to shaft 27 through torquetransfer element 30. Reverse drive is established by applyinglow-and-reverse brake D and friction element B.

For the purpose of illustrating one example of a synchronous ratioupshift for the transmission of FIG. 1, it will be assumed that anupshift will occur between the first ratio and the second ratio. On sucha 1-2 upshift, friction element G starts in the released position beforethe shift and is engaged during the shift while low/reverse frictionelement D starts in the engaged position before the shift and isreleased during the shift. Forward friction element A stays engagedwhile friction element B and overdrive friction element E staydisengaged throughout the shift. More details of this type oftransmission arrangement are found in U.S. Pat. No. 7,216,025, which ishereby incorporated by reference.

FIG. 3 depicts a general process of a synchronous frictionelement-to-friction element upshift event from a low gear configurationto a high gear configuration for the automatic transmission system ofFIG. 1. For example, the process has been described in relation to a 1-2synchronous ratio upshift above wherein friction element C is anoncoming friction element and low/reverse friction element D is anoff-going friction element, but it is not intended to illustrate aspecific control scheme.

The shift event is divided into three phases: a preparatory phase 31, atorque phase 32 and an inertia phase 33. During preparatory phase 31, anon-coming friction element piston is stroked (not shown) to prepare forits engagement. At the same time, off-going friction element controlforce is reduced as shown at 34 as a step toward its release. In thisexample, off-going friction element D still retains enough torquecapacity shown at 35 to keep it from slipping, maintaining transmission2 in the low gear configuration. However, increasing on-coming frictionelement control force shown at 36 reduces net torque flow within gearset21. Thus, the output shaft torque drops significantly during torquephase 32, creating a so-called torque hole 37. A large torque hole canbe perceived by a vehicle occupant as an unpleasant shift shock. Towardthe end of torque phase 32, off-going friction element control force isdropped to zero as shown at 38 while on-coming friction element applyforce continues to rise as shown at 39. Torque phase 32 ends and inertiaphase 33 begins when off-going friction element D starts slipping asshown at 40. During inertia phase 33, off-going friction element slipspeed rises as shown at 41 while on-coming friction element; slip speeddecreases as shown at 42 toward zero at 43. The engine speed andtransmission input speed 44 drops as the planetary gear configurationchanges. During inertia phase 33, output shaft torque indicated byprofile 45 is primarily affected by on-coming friction element C torquecapacity indirectly indicated by force profile 46. When on-comingfriction element C completes engagement or when its slip speed becomeszero at 43, inertia phase 33 ends, completing the shift event.

FIG. 4 shows a general process of a synchronous frictionelement-to-friction element upshift event from the low gearconfiguration to the high gear configuration in which off-going frictionelement D is released prematurely as shown at 51 compared with the caseshown in FIG. 3. When off-going friction element D is released, itbreaks a path between automatic transmission input shaft 18 andautomatic transmission output shaft 24, depicted in FIG. 1, no longertransmitting torque to automatic transmission output shaft at the lowgear ratio. Since on-coming friction element C is yet to carry enoughengagement torque as indicated by a low apply force at 52, automatictransmission output shaft torque drops largely, creating a deep torquehole 53 which can be felt as a shift shock. At the same time, enginespeed or transmission input speed rapidly increases as shown at 54,causing a condition commonly referred to as engine flare. A large levelof engine flare can be audible to a vehicle occupant as unpleasantnoise. Once on-coming friction element C develops sufficient engagementtorque as indicated by a rising control force at 55, automatictransmission input speed comes down and the output torque rapidly movesto a level at 56 that corresponds to on-coming friction element controlforce 55. Under certain conditions, this may lead to a torqueoscillation 57 that can be perceptible to a vehicle occupant asunpleasant shift shock.

FIG. 5 shows a general process of a friction element-to-friction elementupshift event from the low gear configuration to the high gearconfiguration in which off-going friction element release is delayed asshown at 61 compared with the case shown in FIG. 3. Off-going frictionelement D remains engaged even after on-coming friction element Gdevelops a large amount of torque as indicated by a large actual controlforce at 65. Thus, transmission input torque continues to be primarilytransmitted to output shaft 24 at the low gear ratio. However, largeon-coming friction element control force 65 results in a drag torque,lowering automatic transmission output shaft torque, creating a deep andwide torque hole 63. This condition is commonly referred to as a tie-upof gear elements. A severe tie-up can be felt as a shift shock or lossof power by a vehicle occupant.

As illustrated in FIGS. 3, 4, and 5 a missed synchronization ofoff-going friction element release timing with respect to on-comingfriction element torque capacity leads to engine flare or tie-up. Bothconditions lead to varying torque levels and profiles at automatictransmission torque output shaft 24 during shifting. If these conditionsare severe, they result in undesirable driving experience such asinconsistent shift feel or perceptible shift shock. The prior artmethodology attempts to mitigate the level of missed-synchronization byuse of an open loop off-going friction element release control based onspeed signal measurements. It also attempts to achieve a consistenton-coming friction element engagement torque by an open-loop approachduring a torque phase under dynamically-changing shift conditions.

FIG. 6 illustrates a prior art methodology for controlling a frictionelement-to-friction element upshift from a low gear configuration to ahigh gear configuration for automatic transmission 2 in FIG. 1. Theprior art on-coming control depicted in FIG. 6 applies to a conventionaltorque phase control utilized for either a synchronous ornon-synchronous shift. In this example off-going friction element Dremains engaged until the end of torque phase 32. Although the focus isplaced on torque phase control, FIG. 6 depicts the entire shift controlprocess. As shown the shift event is divided into three phases: apreparatory phase 31, a torque phase 32 and an inertia phase 33. Duringpreparatory phase 31, an on-coming friction element piston is stroked(not shown) to prepare for its engagement. At the same time, off-goingfriction element control force is reduced as shown at 34 as a steptoward its release. During torque phase 32 controller 4 commands anon-coming friction element actuator to follow a prescribed on-comingfriction element control force profile 64 through an open-loop basedapproach. Actual on-coming friction element control force 65 may differfrom prescribed profile 64 due to control system variability. Even ifactual control force 65 closely follows prescribed profile 64, on-comingfriction element engagement torque may still vary largely from shift toshift due to the sensitivity of the on-coming friction elementengagement process to engagement conditions such as lubrication oil flowand friction surface temperature. Controller 4 commands enough off-goingelement control force 61 to keep off-going element D from slipping,maintaining the planetary gearset in the low gear configuration untilthe end of torque phase 32. Increasing on-coming friction elementcontrol force 65 or engagement torque reduces net torque flow within thelow-gear configuration. Thus, output shaft torque 66 drops significantlyduring torque phase 32, creating so-called torque hole 63. If thevariability in on-coming friction element engagement torquesignificantly alters a shape and depth of torque hole 63, a vehicleoccupant may experience inconsistent shift feel. Controller 4 reducesoff-going friction element actuator force at 38, following apre-calibrated profile, in order to release it at a pre-determinedtiming 67. The release timing may be based on a commanded value ofon-coming friction element control force 62. Alternatively, off-goingfriction element D is released if controller 4 detects a sign ofsignificant gear tie-up, which may be manifested as a detectable drop ininput shaft speed 44. Inertia phase 33 begins when off-going frictionelement D is released and starts slipping as shown at 67. During inertiaphase 33, off-going friction element slip speed rises as shown at 68while on-coming friction element slip speed decreases toward zero asshown at 69. Transmission input speed 44 drops as the planetary gearconfiguration changes. During inertia phase 33, output shaft torque 66is primarily affected by on-coming friction element torque capacity orcontrol force 65. The shift event completes when the on-coming frictionelement comes into a locked or engaged position with no slip as shown at70.

FIG. 7 illustrates another prior art methodology for controlling torquephase 32 of a synchronous upshift process from the low gearconfiguration to the high gear configuration. In this example,controller 4 allows off-going friction element D to slip during torquephase 32. Although the focus is placed on torque phase control, FIG. 7depicts the entire shift event. During preparatory phase 31, anon-coming friction element piston is stroked to prepare for itsengagement. At the same time, off-going friction element control force86 is reduced as a step toward its slip. During torque phase 32,on-coming friction element control force is raised in a controlledmanner. More specifically, controller 4 commands on-coming frictionelement actuator to follow a prescribed on-coming friction elementcontrol force profile 87 through an open-loop based approach. An actualon-coming friction element control force 88 may differ from thecommanded profile 87 due to control system variability. Even if actualcontrol force 88 closely follows commanded profile 87, on-comingfriction element engagement torque may still vary largely from shift toshift due to the sensitivity of on-coming friction element engagementprocess to engagement conditions such as lubrication oil flow andfriction surface temperature. Increasing on-coming friction elementcontrol force 88 or on-coming friction element engagement torque reducesnet torque flow within the low-gear configuration. This contributes tooutput shaft torque 99 being reduced during torque phase 32, creating aso-called torque hole 85.

If the variability in on-coming friction element engagement torquesignificantly alters the shape and depth of torque hole 85, the vehicleoccupant may experience inconsistent shift feel. A deep torque hole maybe perceived as an unpleasant shift shock. During torque phase 32,off-going friction element control force is reduced as shown at 82 toinduce an incipient slip 83. Controller 4 attempts to maintain off-goingfriction element slip at a target level through a closed-loop controlbased on off-going friction element speed 96 which may be directlymeasured or indirectly derived from speed measurements at pre-determinedlocations. A variability in off-going friction element control force 82of off-going element slip torque may alter the shape and depth of torquehole 85, thus affecting shift feel. If controller 4 inadvertently allowsa sudden increase in off-going friction element slip level, automatictransmission input speed or engine speed 90 may surge momentarily,causing the so-called engine speed flair or engine flair. The engineflair may be perceived by a vehicle occupant as an unpleasant sound.

Controller 4 initiates off-going friction element release process at apredetermined timing shown at which may be based on a commanded value ofon-coming friction element control force 93. Controller 4 lowersoff-going friction element control force, following a pre-calibratedprofile 94. If a release of off-going friction element D is initiatedprematurely before on-coming friction element C develops enough torque,engine speed or input shaft speed may rise rapidly in an uncontrolledmanner. If this engine speed flair 90 is detected, controller 4increases off-going friction element control force to delay off-goingfriction element release process. Alternatively to the pre-determinedoff-going friction element release timing, controller 4 may utilizespeed signals to determine a final off-going friction element releasetiming. When a sign of significant gear tie-up, which may be manifestedas a measurable drop in input shaft speed, is detected, off-goingfriction element D is released following a pre-calibrated force profile.Inertia phase 33 begins when off-going friction element torque capacityor control force drops to non-significant level 95. During inertia phase33, off-going friction element slip speed rises 96 while on-comingfriction element slip speed decreases 97 toward zero. The transmissioninput shaft speed drops as shown at 98 as the planetary gearconfiguration changes. During inertia phase 33, the output shaft torque99 is primarily affected by on-coming friction element torque capacity,which is indicated by its control force 100. When on-coming frictionelement C becomes securely engaged at 101, the shift event completes.

In summary, a prior art methodology, which is based on an open-loopon-coming friction element control during a torque phase, cannot accountfor control system variability and dynamically-changing shift conditionsduring the torque phase, resulting in inconsistent shift feel orunpleasant shift shock. A pre-determined off-going friction elementrelease timing with a pre-calibrated control force profile cannot ensurean optimal timing under dynamically changing shift conditions, resultingin inconsistent shift feel or unpleasant shift shock. The alternativeapproach to gauge off-going friction element release timing based onspeed signals often results in a hunting behavior between gear tie-upand engine flair, leading to inconsistent shift feel. Furthermore,off-going friction element slip control is extremely difficult becauseof its high sensitivity to slip conditions. In addition, a largediscontinuity exists between static and dynamic friction coefficients,introducing a large torque disturbance during an incipient slip control.A failure to achieve a seamless off-going friction element slip controlduring the torque phase leads to undesirable shift shock.

As can be seen from the above discussion the controllability of bothoff-going friction element and on-coming friction element is desirablein order to deliver a consistent and seamless shift quality. The priorart does not have a cost effective design solution to the problem ofdirectly measuring torque passing through either a multiple disc clutchor a band brake and therefore is a need in the art for a transmissioncontrol system that minimizes shift shock during a gear ratio changethat does not rely solely on traditional speed signal measurement or apredetermined open-loop control and instead relies on measuring frictionelement load level in either a multiple plate clutch or a band brake forconsistently controlling its torque level through a closed loopapproach.

SUMMARY OF THE INVENTION

The present invention is directed to a load sensor assembly formeasuring an amount of torque transmitted, through a torque establishingelement of an automatic transmission. The assembly comprises a coremounted on a transmission housing and a load sensor mounted on the coreand positioned against a portion of the torque establishing elementwhereby a portion of the amount of torque transmitted through the torqueestablishing element travels through the load sensor and is measured bythe load sensor assembly. Preferably, a cable is connected to the loadsensor for transmitting a signal representative of the amount of torqueto a transmission controller. A cover or sleeve extends over the coreand the sensor.

In a preferred embodiment, the torque establishing element is a multipledisk friction element including an end plate and a spline connectionbetween the transmission case and the end plate. The connection hasteeth that extend from the transmission case and cooperate with teethextending from the end plate. The load sensor assembly is mounted on thetransmission housing between two spline teeth extending from the endplate and in a location where a spline tooth would normally be located.Preferably, the core is made of metal and the sleeve is made from one ofrubber, plastic and metal. The sensor may have several differentconfigurations. In one configuration, a pin is fixed to the end plateand the load sensor is placed against the pin. In another configuration,the force sensor is a load-resistive elastomer deposited on a thin filmand the core is a tooth of a friction element plate. An example of sucha thin film force sensor can be found in U.S. Pat. No. 6,272,936, whichis incorporated herein by reference. In yet another configuration, thecore is a metal beam securely anchored to the transmission case and theload sensor is a strain sensor that measures an amount of strain on thebeam caused by the torque.

In another embodiment the torque establishing element is a band brakeincluding an anchor bracket and a band brake strap. The core may engagethe strap in many ways. In one configuration, the band brake strap has ablock extending therefrom and the core passes through the transmissionhousing and engages the block. The load sensor is located between thecore and the block. In another configuration, the band brake strap has ahook extending therefrom formed by punching a hole in the strap. Thecore passes through the transmission housing and engages the hook andthe load sensor is located between the core and the hook. In yet anotherconfiguration, the anchor bracket has a pin extending therefrom. Thecore passes through the transmission housing and engages the pin. Theload sensor is located between the core and the pin. Preferably, acushion is located between the load sensor and the cover.

In yet another embodiment, the torque establishing element is a bandbrake including an anchor bracket and a band brake strap while the coreis an anchor pin, which does not necessarily have a cover, mounted inthe transmission case. The anchor pin extends out of the transmissioncase and engages the anchor bracket. The load sensor is mounted betweenthe anchor pin and the transmission case whereby torque is transferredto the band strap, pushes on the anchor pin and is sensed by the loadsensor. Preferably, a cushion is located between the load sensor and theanchor pin. The core includes an anchor pin mounted in the transmissioncase. The core extends out of the transmission case and is connected toan anchor strut which, in turn, engages the anchor bracket. The loadsensor is mounted between the anchor pin and the transmission ease.Torque is transferred to the band strap where it pushes on both theanchor strut and pin, with the torque being sensed by the load sensor.Preferably, the transmission housing includes a hole for supporting theanchor pin. A nut is mounted in one end of the hole and secures theanchor pin to the housing. A plug and a support are located between thenut and the anchor pin. With this arrangement, torque passing throughthe friction elements of a transmission may be directly measured andshift shock and engine flair may be reduced.

Additional objects, features and advantages of the present inventionwill become more readily apparent from the following detaileddescription of preferred embodiments when taken in conjunction with thedrawings, wherein like reference numerals refer to corresponding partsin the several views.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a schematic diagram of a gearing arrangement for an automatictransmission system according to the prior art;

FIG. 2 is a chart showing a clutch and brake friction element engagementand release pattern for establishing each of six forward driving ratiosand a single reverse ratio for the transmission schematicallyillustrated in FIG. 1;

FIG. 3 is a plot of a general process of a Synchronous frictionelement-to-friction element upshift event from a low gear configurationto a high gear configuration for the prior art automatic transmissionsystem of FIG. 1;

FIG. 4 is a plot of the general process of a synchronous frictionelement-to-friction element upshift event from the low gearconfiguration to the high gear configuration in which the off-goingfriction element is released prematurely compared with the case shown inFIG. 3;

FIG. 5 is a plot of the general process of a synchronous frictionelement-to-friction element upshift event from the low gearconfiguration to the high gear configuration in which off-going frictionelement release is delayed compared with the case shown in FIG. 3;

FIG. 6 is plot of a prior art synchronous friction element-to-frictionelement upshift control from a low gear configuration to a high gearconfiguration based on speed measurements of powertrain components forthe automatic transmission system in FIG. 1 wherein an off-goingfriction element remains locked during the torque phase;

FIG. 7 is plot of a prior art synchronous friction element-to-frictionelement upshift control from a low gear configuration to a high gearconfiguration based on speed measurements of powertrain components forthe automatic transmission system in FIG. 1, wherein an off-goingfriction element is slipped during the torque phase;

FIG. 8 is a schematic diagram of a gearing arrangement for an automatictransmission system including load sensor locations in accordance with afirst preferred embodiment of the invention;

FIG. 9 is a plot of a synchronous friction element to friction elementupshift control from a low gear configuration to a high gearconfiguration for the automatic control system in FIG. 8 based on directmeasurements or estimates of torsional load exerted onto an off-goingfriction element in accordance with a preferred embodiment of theinvention;

FIG. 10 is a flow chart showing an on-coming friction element controlmethod in accordance with a preferred embodiment of the invention;

FIG. 11 is a flow chart showing an off-going element release controlmethod in accordance with a preferred embodiment of the invention;

FIG. 12 is a plot of the process used to determine an ideal releasetiming of the off-going friction element in accordance with firstpreferred embodiment of the invention;

FIG. 13 is a flow chart showing a shift control method in accordancewith a preferred embodiment of the invention;

FIG. 14 is a plot of a synchronous friction element-to-friction elementupshift from a low gear configuration to a high gear configuration forthe automatic transmission control system in FIG. 8 based on the directmeasurements or estimates of torsional load exerted onto an off-goingfriction element and an on-coming element in accordance with anotherpreferred embodiment of the invention;

FIG. 15 is a flow chart showing an on-coming friction element shiftcontrol method in accordance with another preferred embodiment of theinvention;

FIG. 16A depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed between two teeth of aendplate of a friction element for measuring a relative load level onthe friction element;

FIG. 16B depicts the load sensor assembly of FIG. 16A installed in atransmission case;

FIG. 17A depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention placed against a pin extendingfrom an endplate of a friction element for measuring a relative loadlevel on the off-going friction element;

FIG. 17B depicts the load sensor assembly of FIG. 17A installed in atransmission case;

FIG. 18 depicts a load sensor in accordance with another preferredembodiment of the invention formed of a thin film-type load sensor andattached to a tooth for measuring a relative load level on the off-goingfriction element;

FIG. 19 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention formed of a metal beam formeasuring a relative load level on the off-going friction element;

FIG. 20 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement;

FIGS. 21A-21C depict a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement;

FIGS. 22A and 22B depict a load sensor assembly in accordance withanother preferred embodiment of the invention installed on a band braketype friction element for measuring a relative load level on thefriction element;

FIG. 23 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement;

FIG. 24 depicts a chart in accordance with another preferred embodimentof the invention;

FIG. 25 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement;

FIG. 26 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement;

FIG. 27 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement;

FIG. 28 depicts a load sensor assembly in accordance with anotherpreferred embodiment of the invention installed on a band brake typefriction element for measuring a relative load level on the frictionelement; and

FIG. 29 depicts a chart in accordance with another preferred, embodimentof the invention.

DESCRIPTION OF PREFERRED EMBODIMENTS

With initial reference to FIG. 8, there is shown an automotivetransmission employing the invention. As this automatic transmissionarrangement is similar to the one schematically illustrated in FIG. 1all the same parts have been indicated with corresponding referencenumbers and therefore a duplicate discussion of these parts will not bemade here. Instead, of particular importance is the addition of a torquesensor 120 located in friction element C, a load sensor 130 located infriction element D, and a torque sensor 131 located in transmissionoutput shaft 24, all connected to controller 4 for controlling variousfunctions of transmission 2 as will be more fully discussed below.

FIG. 9 shows a torque phase control method according to a preferredembodiment of the invention for a synchronous frictionelement-to-friction element upshift from a low gear configuration to ahigh gear configuration for the automatic transmission system in FIG. 8.The on-coming friction element control method illustrated here is alsoapplicable to non-synchronous shift control. The shift event is dividedinto 3 phases: preparatory phase 31, torque phase 32 and inertia phase33. During preparatory phase 31, an on-coming friction element piston isstroked to prepare for its engagement. At the same time, off-goingfriction element control force or its torque capacity is reduced asshown at 404 as a step toward its release. During torque phase 32,on-coming friction element control force is raised in a controlledmanner as shown at 405. More specifically, controller 4 commandson-coming friction element actuator to follow a target on-comingfriction element engagement torque profile 406 through a closed-loopcontrol directly based on the measurements of on-coming friction elementengagement torque 407 during torque phase 32. On-coming friction elementtorque 407 may be directly measured using a load sensor according tothis invention as more fully described below. On-coming friction elementengagement torque directly affects transmission output torque that istransmitted to the vehicle wheels. This torque-based close-loop controleliminates or significantly reduces the undesirable effects of on-comingfriction element engagement torque sensitivity to hardware variabilityand shift conditions, achieving a consistent shift feel, regardless ofshift conditions.

Alternatively to the direct measurements, on-coming friction elementtorque can be determined from the measurements of transmission outputshaft torque using torque sensor 131 depicted in FIG. 8. Mathematically,on-coming friction element torque T_(OCE) can be described as a functionof measured output shaft torque T_(OS) as:T _(OCE)(t)=G _(OCE) T _(OS)(t)  Eq. (1)Where G_(OCE) can be readily obtained based on a given gear setgeometry.

Yet alternatively, on-coming friction element torque T_(OCE) can beestimated through the following Eq. (2), based on a slight change intransmission component speeds ω_(i) at pre-determined locations (i=1, 2,. . . , n),T _(OCE)(t)=F _(trans)(ω_(i) ,t)  Eq. (2)where t indicates time and F_(trans) represents a mathematicaldescription of a transmission system. More specifically, as on-comingfriction element engagement torque rises 407, torque levels transmittedthrough various transmission components change. This creates small, butdetectable changes in ω_(i). A transmission model, F_(trans), can bereadily derived to estimate on-coming friction element engagement torquewhen off-going friction element remains locked during torque phase 32.

Controller 4 commands enough off-going friction element control force408 to keep it from slipping, maintaining the planetary gearset in thelow gear configuration during torque phase 32. As on-coming frictionelement engagement torque 407 increases, a reaction torque goes againsta component that is grounded to a transmission case. More specifically,in this case, torque transmitted through off-going friction element ortorsional load 409 exerted onto off-going friction element D decreasesproportionally. Off-going friction element load level 409 can bedirectly monitored using a torque sensor such as is more fully discussedbelow. Alternatively, off-going friction element load level T_(OGE) 409can be calculated from measured or estimated on-coming friction elementengagement torque T_(OCE) 407 when off-going friction element remainslocked during torque phase 32 according to:T _(OGE)(t)=F _(OCE/OGE)(T _(OCE)(t))  Eq. (3)where F_(OCE/OGE) represents a torque ratio between on-coming frictionelement C and off-going friction element D at the low gear configurationand can be obtained based on gear set geometry. According to thisinvention, off-going friction element D is released at an ideal timingwhen torque load exerted onto off-going friction element D becomes zeroor a near-zero level. Transmission controller 4 initiates a releaseprocess of off-going friction element D as shown at 410 as off-goingfriction element load 409 approaches zero at 411. Off-going frictionelement torque is dropped quickly as shown at 412 with no slip control.Since no off-going friction element slip control is involved, the methodis insensitive to off-going friction element break-away frictioncoefficient variability. In addition, the quick release of off-goingfriction element D shown at 412 induces little disruption in outputshaft torque at 413 because off-going friction element load level isnear zero as shown at 411 at the moment of release. Off-going frictionelement D starts slipping 411 once its control force reaches anon-significant level. During inertia phase 33, a conventional controlapproach may be utilized based on on-coming friction element slipmeasurements. Off-going friction element slip speed increases as shownat 415 while on-coming friction element slip speed decreases as shown at416. The transmission input speed drops as shown at 417 as the planetarygear configuration changes. During inertia phase 33, output shaft torque418 is primarily affected by on-coming friction element torque level419. Alternatively to the conventional control, a closed loop controlthat is based on measured or estimated on-coming friction element torquemay continue to be employed. When on-coming friction element C completesengagement or when its slip speed becomes zero as shown at 420, theshift event completes.

FIG. 10 shows a flow chart of closed-loop on-coming friction elementengagement torque control during the torque phase depicted in FIG. 9.Step 430 is the beginning of torque phase 32. Controller 4 chooses adesired on-coming element torque at step 431 and measures or estimatesan actual torque at step 432. At step 433, the on-coming frictionelement actuator is then adjusted by controller 4 based on thedifference between the measured/estimated torque level and the actualtorque level. At step 434, controller 4 determines if torque phase hasended and if so controller 4 starts inertia phase 33 at 436.

FIG. 11 shows a flow chart of an off-going friction element torquecontrol process during torque phase 32 depicted in FIG. 9. The processstarts at step 440 at the beginning of torque phase 32. A loadtransmitted through locked off-going friction element D is directlymeasured or estimated at step 441. At step 442, when its load leveldrops below a predetermined level, off-going friction element D ispromptly released at step 444. The control process ends at step 445 atthe end of torque phase 32.

Alternatively to the measurements or estimates of absolute load levels,FIG. 12 illustrates the process to determine the ideal release timing ofoff-going friction element D based on relative load measurements orestimates according to this invention. FIG. 12 depicts an actual loadprofile 451 exerted on off-going friction element D and a relative loadprofile L(t) 452 measured by torque sensor 130 during the upshift eventin FIG. 9. The preferred embodiment requires only relative load profileL(t) 452. Relative load profile L(t) 452 is preferably constructed fromuncalibrated sensor output that reflects actual load profile 451, butnot its absolute levels. This feature eliminates the need of a fullsensor calibration across the entire load range. It also makes thepreferred embodiment insensitive to sensor output drift over time.However, the preferred embodiment relies on knowledge of sensormeasurement L₀ 453 which corresponds to zero off-going friction elementload level 454. Sensor measurement L₀ 453 can be readily identified, asoften as required, by sampling sensor output while vehicle transmission2 is in a neutral or a similar condition where no load is exerted ontooff-going friction element D. Transmission controller 4 collectsrelative load data 455 during torque phase 32 to dynamically constructrelative load profile L(t) 452. Then, controller 4 extrapolates L(t) topredict t₀ 457 where L(t₀)=L₀. Once t₀ 457 is obtained in advance,controller 4 predicts when to initiate an off-going friction elementrelease process. Specifically controller 4 starts the release process ata time equal to t₀-Δt shown, at 458, where Δt is the time required toquickly drop off-going friction element control force to zero. In thisway, off-going friction element D starts slipping at or near idealtiming t₀ 457 when the actual off-going friction element load level isat or close to zero as shown by reference numeral 454.

FIG. 13 presents a flow chart of the new upshift control methodaccording to this invention. During preparatory phase 31 at step 461 ofa synchronous upshift event, off-going friction element torque capacityor apply force is reduced to a holding level without allowing any slipat step 462 while on-coming friction element piston is stroked at step463. During torque phase 32, transmission controller 4 measures at step465 a relative load level exerted onto off-going friction element D at apre-specified sampling frequency using torque sensor 130 describedfurther below. Controller 4 repeats this measurement step 465 untilenough data points are collected at step 466 for dynamicallyconstructing a relative load profile at step 467 that shows load as afunction of time L(t). Once relative load profile L(t) is obtained,controller 4 predicts the ideal off-going friction element releasetiming t₀ at step 468 so that L(t₀)=L₀ where L₀ corresponds to asubstantially zero load level on off-going friction element D.Controller 4 initiates an off-going friction element release process att₀-Δt as shown as step 469 where Δt is a pre-specified time required toquickly drop off-going friction element apply force to zero.Alternatively, controller 4 may initiate the off-going friction elementrelease process at t_(thres) such that L(t_(thres))=L_(thres) whereL_(thres) is a predetermined threshold. No slip control is required foroff-going friction element D during torque phase 32. Inertia phase 33starts when off-going friction element D is released. The controlmethodology illustrated in FIG. 10 is preferably applied to on-comingfriction element C during torque phase 32. A conventional on-comingfriction element control may be applied during inertia phase 33 based onspeed signals. When on-coming friction element C becomes securelyengaged at step 473, the shift event completes at step 474.

FIG. 14 illustrates another preferred embodiment of the inventionrelating to a transmission system with an on-coming friction elementactuator which may not have a sufficient control bandwidth compared witha sampling time of load measurements. At the beginning of torque phase32, a transmission controller raises on-coming friction element actuatorforce based on a pre-calibrated slope 480 over a time interval Δtbetween t₀ and t₁ as shown at interval 481. During interval 481,on-coming friction element load is either measured or estimated with asampling time finer than Δt to construct an engagement torque profile482. If the measured or estimated torque profile 482 indicates a slowrise compared with a target torque profile 483, controller 4 increases aslope of commanded on-coming friction element control force for a nextinterval 485 between t₁ and t₂. On the other hand, if the actual torqueis rising faster than a target profile, controller 4 reduces a slope ofcommanded on-coming friction element control force. For example, duringinterval 485 between t₁ and t₂, on-coming friction element load iseither measured or estimated with a sampling time finer than Δt toconstruct an engagement torque profile 486. The measured or estimatedslope 486 of the engagement torque is compared against a target profile487 to determine a slope 488 of commanded force profile for thefollowing control interval. This process is repeated until the end oftorque phase 32. The off-going friction element release control remainsthe same as that shown in FIG. 9.

FIG. 15 shows a flow chart of alternative closed-loop on-coming frictionelement engagement torque control during torque, phase depicted in FIG.14. The start of torque phase 32 is shown at step 520. Following path521, the off-coming friction element torque is measured or estimated atstep 522 and torque profile 482 is created therefrom at step 523. Themethod may have to go through several iterations as shown by decisionblock 524 and return loop 525. Torque slope profile 486 or an averagederivative of torque profile 482 is calculated at 526 and while adesired target slope profile 487 is calculated at 527 and compared withtorque slope profile 486 at 528. The actuator force Slope is increased529 or decreased 530 and the process continues 531, 532 until the end oftorque phase 32. The process then proceeds to inertia phase 33 at 533.

While the shift control has been discussed above, attention is nowdirected to the structure of the various load sensor assemblies. FIGS.16A, 16B, 17A, 17B, 18 and 19 depict several preferred embodiments ofload sensor assemblies for measuring a relative load level exerted onoff-going friction element D or on-coming element C according topreferred embodiments of the invention. FIG. 16A shows a cross-sectionalview of a load sensor assembly 601 design according to a preferredembodiment. In FIG. 16A, sensor assembly 601 is installed between twoteeth 602, 603 of an end plate 604 of off-going friction element D.Assembly 601 includes a core 605, a load sensor 606 and a sleeve 607.Core 605 is preferably made from a metal, such as steel or aluminum, andis securely grounded to a transmission case 608 through anchor bolts609. Load sensor 606 is preferably a film-type sensor constructed with apressure-resistive material. Sensor 606 generates an electrical signalthat corresponds to a relative level of loading force 610. Sleeve 607,which protects sensor 606, is preferably made from rubber, plastic ormetal. While cover 607 is referred to as either a sleeve or a cover, itis to be understood that the terms are interchangeable. FIG. 16Billustrates an installation of sensor assembly 601 in transmission case608. Sensor assembly 601 is securely positioned in a location where aspline tooth is normally located otherwise. When off-going frictionelement plates are installed, end plate 604 fits snugly around sensorassembly 601, providing a preload to sensor 606. That is, sensor 606preferably indicates non-zero output L₀ even when no load is exerted onoff-going friction element D or its end plate 604. When the torque loadis exerted as shown by arrow 610 during a shift event, the output fromsensor 601 provides a relative measure of the load on off-going frictionelement D. When this embodiment is employed to measure relative loadexerted onto an off-going friction element such as when torque sensor130 is used to measure the load on friction element D, it is readilyunderstood that optimal friction element release timing is identifiedwhen the sensor output level approaches to L₀ corresponding to zero loadlevel.

FIGS. 17A and 17B depict another sensor assembly 611 which has a similarstructure to assembly 601 in FIG. 16A. Assembly 611 includes a groundedcore 612, a force sensor 613 and a sleeve 614. However, as illustratedin FIG. 17A, assembly 611 is placed against a pin 615 that is fixed toan end plate 616 of off-going friction element D. Sensor 613 ispreloaded against pin 615, providing non-zero output in the absence oftorque load on off-going friction element end plate 616 (FIG. 17B). Whena torque load is exerted on off-going friction element D, pin 615 ispressed with a force 617 against sensor 613 across sleeve 614. Thisenables sensor 613 to provide the relative measure of torque load onoff-going friction element D. FIG. 17B shows a view of sensor assembly611 and off-going friction element end plate 616 with pin 615 in atransmission case 618.

FIG. 18 shows another potential embodiment of this invention wherein athin film-type force sensor 621 is directly attached to a tooth 622 of afriction element plate 623, covered with a protective sleeve 624. Sleeve624 is preferably made from rubber, plastic or metal. When plate 623 isinstalled into a transmission case 625, sensor 621 directly measurescontact load 626 between friction element tooth 622 and a spline 627through sleeve layer 624, providing a relative measure of the loadexerted onto off-going friction element D.

FIG. 19 shows another preferred embodiment of the invention wherein ametal beam 631, which is securely anchored to a transmission case 632,is installed and positioned between two teeth 633, 634 of an off-goingfriction element plate 635. As a load level 636 exerted on plate 635varies, a strain level of beam 631 changes. The level of the strain isdetected through a strain sensor 637, providing a relative measure oftorque load exerted on off-going friction element D. Optionally, a covermay be added to protect strain sensor 637.

FIGS. 20, 21A, 21B, 21C, 22A, 22B and 23-29 show various preferredembodiments of the invention relating to directly measuring torque in afriction element. More specifically, FIG. 20 shows a partial view of aband brake system 700 with a load sensing assembly 731. Brake system 700includes an anchor end of a band strap 732, a pin or a hook 733, and ananchor bracket 734. Band strap 732 is preferably either a single-wrap ordouble-wrap type. Load sensor assembly 731 includes an assembly core735, a load sensing unit 736 and a protective sleeve or cover 737.Assembly core 735 is made of a metal and securely mounted to atransmission case 738 with a bolt 739 or any other means. Cover 737 maybe made of metal, rubber, plastic or any other materials. Cover 737protects sensor unit 736 from direct contact with pin or hook 733 forreduced sensor material wear. Cover 737 may be made of athermally-insulated material to protect sensor 736 from heat. Cover 737also acts as a protective shield against any other hostile conditionsthat include electro-chemical reaction with transmission oil. Loadsensing unit 736, which may be a pressure resistive film-type, ispositioned between core 735 and cover 737. The tip of sensor 736 ispositioned against pin 733 across cover 737. When a band engagement iscommanded, strap 732 is pulled by a hydraulic servo (which is describedbelow) in the direction shown with an arrow 740. Band strap 732stretches slightly, pushing pin or hook 733 against load sensor 736.Load sensor 736 generates an electrical signal according to a magnitudeof the contact force. That is, sensor 736 provides a relative measure ofband tension at the location of pin 733. The electrical signal istransmitted to a data acquisition unit (not shown) and then tocontroller 4 through an electrical cable 741.

FIGS. 21A, 21B and 21C depict band strap designs in detail. In FIG. 21A,a band strap 732 has a part punched out and bent to form a pin or a hook753 and a hole 752. Hole 752 also acts as an oil drain during bandengagement. In FIG. 21B, a small pin or a block 754 is riveted, screwedor welded to strap 732. Alternatively, a pin or a hook 755 can be formedas apart of an anchor bracket 734 as shown in FIG. 21C. A pin 755 isattached to a band anchor bracket 734 instead of a strap 732. Sensorassembly 731 is positioned against the pin 755. Since bracket 732 isstiffer than the strap 732, its strain is smaller under loadedconditions during both holding and engagement. Thus, a level of forceexerted onto a load sensor 736 through a micro displacement of pin 755is reduced significantly. The lower stress level improves the life ofthe sensor assembly 731 while enabling the use of a sensor 736 rated fora lower maximum force.

FIG. 22A illustrates sensor functions during a band engagement process.When the engagement is initiated, transmission controller 4 sends anelectrical signal I(t) to raise and regulate a hydraulic force 761applied to a servo piston 762. As servo piston 762 is stroked, a servorod 763 pulls one end 764 of band strap 732. Tension around strap 732builds up, squeezing out lubrication oil 766 from a band-drum interface.During the engagement, brake torque from strap 732 to a drum 767 ispartly transmitted through viscous shear across oil 766. The braketorque is transmitted through a mechanical frictional force once strap732 makes physical contact with drum 767. According to a conventionalanalysis, the relationships between engagement torque T_(eng), bandtension at a pin F_(pin) 733 and band tension at a servo F_(servo) 769can be written as follows, assuming a Coulomb friction model as aprimary torque transfer mechanism between band strap 732 and drum 767:T _(eng) =F _(servo) R(e ^(μβ)−1)  Eq. (4)F _(pin) =F _(servo) e ^(μβ)  Eq. (5)where R=drum radius, m=a Coulomb friction coefficient, b=a band wrapangle 770 assuming that pin 733 is positioned sufficiently close to ananchor 734. Drum 767 rotates in the same direction 772 as the hydraulicforce 761. Substituting Eq. (5) into Eq. (4) yields:

$\begin{matrix}{T_{eng} = {{F_{pin}{R\left( {1 - {\mathbb{e}}^{- {\mu\beta}}} \right)}\mspace{14mu}{or}\mspace{14mu} F_{pin}} = \frac{T_{eng}}{R\left( {1 - {\mathbb{e}}^{- {\mu\beta}}} \right)}}} & {{Eq}.\mspace{14mu}(6)}\end{matrix}$Since the electrical output signal S_(pin) from the sensor isapproximately linear with band tension F_(pin):S _(pin) =kF _(pin)  Eq. (7)where k is a proportional constant. Substituting Eq. (7) into Eq. (6)yields:

$\begin{matrix}{S_{pin} = {{\frac{k}{R\left( {1 - {\mathbb{e}}^{- {\mu\beta}}} \right)}T_{eng}} = {{k^{\prime}T_{eng}\mspace{14mu}{or}\mspace{14mu}\frac{\mathbb{d}S_{pin}}{\mathbb{d}t}} = {k^{\prime}\frac{\mathbb{d}T_{eng}}{\mathbb{d}t}}}}} & {{Eq}.\mspace{14mu}(8)} \\{{{where}\mspace{14mu} k^{\prime}} = \frac{k}{R\left( {1 - {\mathbb{e}}^{- {\mu\beta}}} \right)}} & {{Eq}.\mspace{14mu}(9)}\end{matrix}$According to Eq. (8), the sensor output S_(pin) provides a relativemeasure of band brake engagement torque T_(eng).

This embodiment provides a relative measure of T_(eng) and itsderivative (dT_(eng)/dt) that enables a closed loop control of on-comingfriction element engagement process during torque phase 32. Itsignificantly improves band engagement control, mitigating a sudden riseof band brake torque known as “grabbing” behaviors. Alternatively, thesensor signals may be utilized to adaptively optimize open-loopcalibration parameters such as a rate of pressure rise as a function ofoil temperature in order to achieve a consistent (dT_(eng)/dt). Thesimilar analysis can be applied to the so-called “de-energized” bandengagement where the drum spins in the opposite direction of the servo.

FIG. 22B illustrates sensor functions while band strap 732 is securelyengaged around drum 767 under a holding condition without any slippage.In this case, the band tension F_(pin) at pin 733 reflects both thelevel of the band tension F_(servo) 784 at the servo and the level oftorque load T_(load) 785 exerted onto band 732 and drum 767 from theadjoining components (not shown). It is important that one shouldclearly differentiate T_(load) from T_(eng) which is brake torqueexerted from the band to the drum under slipping conditions.

According to a conventional analysis, the relationships between F_(pin),F_(servo) and T_(load) can be algebraically written as:

$\begin{matrix}{F_{pin} = {{F_{servo} + {\frac{T_{load}}{R}\mspace{14mu}{or}\mspace{14mu} T_{load}}} = {R\left( {F_{pm} - F_{servo}} \right)}}} & {{Eq}.\mspace{14mu}(10)}\end{matrix}$Substituting Eq. (10) into Eq. (7), the sensor output S_(pin) can bedescribed as a function of F_(servo) and T_(load) as:

$\begin{matrix}{S_{pin} = {{kF}_{pin} = {{kF}_{servo} + {\frac{k}{R}T_{load}}}}} & {{Eq}.\mspace{14mu}(11)}\end{matrix}$Note that F_(servo) is a function of an electrical signal I commanded toa hydraulic control system from a transmission controller. That is:F _(servo) =F _(servo)(I)  Eq. (12)Substituting Eq. (12) into Eq. (11) results in:

$\begin{matrix}{S_{pin} = {{kF}_{pin} = {{{kF}_{servo}(I)} + {\frac{k}{R}T_{load}}}}} & {{Eq}.\mspace{14mu}(13)}\end{matrix}$In the absence of T_(load), Eq. (13) becomes:S _(pin) =kF _(servo)(I)=S _(pin) ^(noload)(I)  Eq. (14)where S_(pin) ^(noload) is defined as the sensor output measured underno load condition for a given level of I. In practice S_(pin) ^(noload)can be readily obtained, as required, by sweeping the servo actuatorwith a varying level of I while a vehicle is in a stationary condition.Substituting Eq. (14) into Eq. (13) yields.

$\begin{matrix}{{S_{pin} - {S_{pin}^{noload}(I)}} = {\frac{k}{R}T_{load}}} & {{Eq}.\mspace{14mu}(15)}\end{matrix}$Thus, S_(pin)−S_(pin) ^(noload)(I) provides a relative measure of torqueload T_(load) for a given electrical input I. The optimal timing torelease off-going friction element during a synchronous shift is whenthe load exerted onto off-going friction element or T_(load) becomeszero. This can be readily determined by sampling S_(pin) and evaluatingS_(pin)−S_(pin) ^(noload)(I) for a given electrical signal I. The use ofthe load sensor assembly according to this embodiment significantlyimproves band release controllability during a synchronous shift underall the operating conditions.

FIG. 23 shows a cross-sectional view of another sensor assembly 811including a cushion element 812 inserted between a load sensor 813 and apin or a block 814 that is attached to a band strap or an anchorbracket. Cushion element 812 is preferably made of a rubber.Alternatively, cushion element 812 may be made of a metal in the form ofa spring such as a disk spring or a conical spring. A protective cover815 is preferably positioned between cushion element 812 and block 814.Cover 815 is readily slidable at a nominal force under loadedconditions. The loading force is transmitted from block 814 to loadsensor 813 by deformation of cushion element 812. Accordingly, cushionelement stiffness is used to specify a force range at sensor 813 for agiven range of loading force at block 814. The force transmitted to loadsensor 813 becomes limited once the cushion element surface becomesflush with surface 817 of the assembly core. This non-linearcharacteristic indicated at 818 enables high resolution forcemeasurement for a targeted load range 819 as shown in FIG. 24 whileprotecting sensor 813 from excessive loading.

FIG. 25 shows an alternative embodiment of this invention. In thisdesign, a load sensor 821 is placed at the bottom of a band anchor pin822 inside a transmission case 823. Electrical cable 824 attached tosensor 821 is routed outside through case 823. The tip of pin 822 isinserted into an anchor bracket 826, which is attached to band strap825. When the band brake system is actuated, strap 825 is hydraulicallyor mechanically tightened around a drum such that anchor bracket 826pulls pin 822 in the direction of anchor load 828 as represented by anarrow. Accordingly, load sensor 821 directly measures an anchor load 828exerted onto pin 822 from the anchor bracket 826. A cushion element 831is preferably placed between the bottom of an anchor pin 822 and loadsensor 821. Note that the sensing area of sensor 821 is smaller than thesurface area of cushion element 831. The anchor load supported by pin822 is distributed over the surface of cushion element 831. Accordingly,only part of the anchor load is transmitted to load sensor 821. Thisenables the use of a sensor rated for a lower maximum force.

In FIG. 26, a strut 841 is inserted between an anchor bracket 826 and ananchor pin 843. Strut 841 enables the flexible placement of anchor pin843 with respect to band strap 825 and transmission case 823. Also, anangle 845 between strut 841 and pin 843 may be adjusted to optimize alevel of the axial loading that bracket 876 exerts onto pin 843 throughstrut 841. Cushion element 831 and the reduced axial loading allow theuse of a sensor 821 rated for a lower maximum force. Alternatively,angle 845 may be adjusted to reduce the side loading onto pin 843 tominimize sensor output hysteresis caused by sticky pin displacementunder the loaded conditions.

The embodiment of the invention in FIG. 27 shares many of the samefeatures described in connection with the embodiment in FIG. 26. First,anchor pin 853 is inserted into an unthreaded hole 852 insidetransmission case 823. Its large head 854 prevents pin 853 from fallingthrough hole 852. A cushion element 836 and a load sensor 821 are placedagainst pin head 854. Cushion element 836 may be made of a rubber andact as a seal to protect the sensor 821 from transmission oil. Behindsensor 821 and cushion element 836 is a sensor support dish 857, whichmay be made of a metal. Sensor support dish 857 is backed by a largeplug 858 inserted into a threaded hole 859. The position of plug 858 maybe adjusted and locked with a nut 860 in order to set anchor pin 853 toa desirable position with respect to anchor bracket 826 and strut 841.

The embodiment of the invention shown in FIG. 28 shares features withthe embodiment for FIG. 27. Specifically, a load sensor 821 is placedbehind a cushion element 872 inside support dish 874 with a raisedretaining wall 873. Cushion element 872 is preferably made of rubber.Alternatively, cushion element 872 may be made of metal in the form of aspring such as a disk or a conical spring. Under a no load condition,the surface of cushion element 872 is in contact with that of a pin 875,while the end of retaining wall 873 is away from the surface of pin 875.When the anchor load is below a predetermined level, the entire load istransmitted to sensor 821 through the elastic deformation of cushionelement 872. As the anchor load increases, cushion element 872 becomescompressed. Once the surface level of cushion element 872 becomes flushwith the end of retaining wall 873, retaining wall 873 starts supportingthe load exerted on pin 875, limiting the load on sensor 821.

As shown in FIG. 29, cushion element stiffness determines where thesensor output starts leveling off at 876. This embodiment of theinvention enables the sensor performance to be targeted for a specificload range, maximizing a measurement resolution 877. In addition, sensoroutput voltage at limiting load level 876 and at zero load level 878 canbe used to auto-calibrate sensor 821 for enabling absolute loadmeasurements. That is when the sensor output reaches its maximumplateau, a transfer function between sensor output voltage and loadlevel can be mapped based on two point calibration. This feature isextremely useful, especially if sensor characteristics drift over timeor vary under different operating conditions. This load-limiting featurealso protects the sensor from overloading, preventing its failure.

Based on the above, it should be readily apparent that the presentinvention provides numerous advantages over prior friction elementcontrol during a torque phase of gear-ratio changing. The preferredembodiments provide a consistent output shaft torque profile for apowertrain system with a step-ratio automatic transmission system duringa synchronous friction element-to-friction element upshift, whichreduces shift shock. Also, there is a significant reduction in shiftfeel variability for a powertrain system with a step-ratio automatictransmission system during a synchronous friction element-to-frictionelement upshift. The preferred embodiments of the invention permit theuse of either absolute or relative load levels which are directlymeasured or estimated. The use of a relative load profile, instead ofabsolute load levels, eliminates the need of full-sensor calibration,while the use of a relative load profile only requires one point sensorcalibration that corresponds to zero load level and improves robustnessagainst sensor drift over time. The preferred embodiments also providefor reduced output shaft torque oscillation at the beginning of theinertia phase due to the release of the off-going friction element at ornear the ideal release timing where its load level is zero or close tozero and robustness against the variability of off-going frictionelement breakaway friction coefficient by means of a quick release ofthe off-going friction element at the ideal synchronization timing.

Further advantages include a consistent output shaft torque profile andsignificant reduction in shift feel variability for a powertrain systemwith a step-ratio system during a torque phase of a synchronous frictionelement-to-friction element upshift and during a torque phase of anon-synchronous upshift with an overrunning coupling element. Further,the system provides robustness against the variability of off-goingfriction element breakaway friction coefficient by means of a quickrelease of an off-going friction element at an ideal synchronizationtiming during a synchronous shift and against the variability of afriction element actuation system for both synchronous andnon-synchronous shifts.

A clutch load sensor assembly provides a relative measure of torque loadexerted to the clutch while it is engaged. A band brake load sensorassembly provides a relative measure of engagement torque (brake torque)and its derivative during an engagement process while a band slipsagainst a drum and a relative measure of torque load exerted onto a bandand a drum while the band is securely engaged to the drum withoutslippage. Sensor output may be calibrated with respect to a commandsignal to a band servo actuator while torque load is zero. Use of aprotective cover in the sensor assembly prevents a direct contactbetween a load sensing material and the pin for reduced sensor materialwear; and shields the sensor from hostile conditions that include heatand electro-chemical interaction, such as with transmission oil.

Although described with reference to preferred embodiments of theinvention, it should be understood that various changes and/ormodifications can be made to the invention without departing from thespirit thereof. For example, the invention could be extended to adoubles-wrap band brake system. In general, the invention is onlyintended to be limited by the scope of the following claims.

1. In a multiple-ratio automatic transmission for an automotive vehiclepowertrain, the automatic transmission including a controller, atransmission housing, an input torque source, a torque output member,gearing defining multiple torque flow paths from the input torque sourceto the torque output member and a pressure actuated torque establishingelement for establishing a gear configuration with a low speed ratio ora gear configuration with a higher speed ratio during a ratio upshiftevent having a preparatory phase, a torque phase and an inertia phase, aload sensor assembly for measuring an amount of torque transmittedthrough the torque establishing element comprising: a core mounted onthe transmission housing; and a load sensor mounted on the core andpositioned against a portion of the torque establishing element wherebya portion of the amount of torque transmitted through the torqueestablishing element travels through the load sensor and is measured bythe load sensor assembly, wherein a signal representative of the amountof torque transmitted through the torque establishing element is sent tothe controller from the load sensor.
 2. The load sensor assembly ofclaim 1, further comprising a cover extending over the core and the loadsensor.
 3. The load sensor assembly of claim 2, wherein the torqueestablishing element is a multiple disk friction element including anend plate, a spline connection between transmission housing and the endplate, said connection having spline teeth that extend from saidtransmission housing and interengage with spline teeth that extend fromthe end plate.
 4. The load sensor assembly of claim 3, wherein the loadsensor assembly is mounted on the transmission housing between two ofthe spline teeth extending from the end plate and in a location where aspline tooth would normally be located.
 5. The load sensor assembly ofclaim 3, wherein the core is made of metal and includes a sleeve made ofa material selected from the group consisting of rubber, plastic andmetal.
 6. The load sensor assembly of claim 3, further comprising a pinfixed to the end plate, wherein the load sensor is positioned againstthe pin.
 7. The load sensor assembly of claim 3, wherein the load sensoris a thin load-resistive film and the core is a spline tooth extendingfrom a friction element plate.
 8. The load sensor assembly of claim 3,wherein the core is a metal beam securely anchored to the transmissionhousing and the load sensor is a strain sensor that measures an amountof strain on the core caused by the amount of torque transmitted throughthe torque establishing element.
 9. The load sensor assembly of claim 2,wherein the torque establishing element is a band brake including ananchor bracket and a band brake strap.
 10. The load sensor assembly ofclaim 9, wherein the band brake strap has a block extending therefrom,the core passes through the transmission housing and engages the block,and the load sensor is located between the core and the block.
 11. Theload sensor assembly of claim 9, wherein the band brake strap has a hookextending therefrom, the core passes through the transmission housingand engages the hook, and the load sensor is located between the coreand the hook.
 12. The load sensor assembly of claim 9, wherein theanchor bracket has a pin extending therefrom, the core passes throughthe transmission housing and engages the pin, and the load sensor islocated between the core and the pin.
 13. The load sensor assembly ofclaim 9, further comprising a cushion or spring element located betweenthe load sensor and the cover.
 14. The load sensor assembly of claim 1,wherein the torque establishing element is a band brake including ananchor bracket and a band strap.
 15. The load sensor assembly of claim14, wherein the core is an anchor pin mounted in the transmissionhousing, extending out of the transmission housing and engaging theanchor bracket, and wherein the load sensor is mounted between theanchor pin and the transmission housing, whereby the amount of torquetransmitted through the torque establishing element is transferred tothe band strap, pushes on the anchor pin and is sensed by the loadsensor.
 16. The load sensor assembly of claim 15, further comprising acushion element located between the load sensor and the anchor pin. 17.The load sensor assembly of claim 16, further comprising a support dishwith a retaining wall surrounding a sensor core, wherein the cushionelement touches the anchor pin while the retaining wall of the supportdish is spaced from the anchor pin, under no load condition.
 18. Theload sensor assembly of claim 17, wherein under a load condition thecushion element is compressed and the retaining wall of the support dishcontacts the anchor pin to limit the load exerted on the sensor core.19. The load sensor assembly of claim 15, wherein an anchor pin extendsout of the transmission housing and is connected to an anchor strutengaging the anchor bracket, and wherein the load sensor is mountedbetween the anchor pin and the transmission housing, whereby the amountof torque transmitted through the torque establishing element istransferred to the band strap, which pushes on the anchor strut and theanchor pin, and is sensed by the load sensor.
 20. The load sensorassembly of claim 19, wherein the transmission housing includes a holefor supporting the anchor pin and further comprising a nut for fillingone end of the hole and securing the anchor pin to the housing.
 21. Theload sensor assembly of claim 20, further comprising a plug and asupport located between the nut and the anchor pin.
 22. A load sensorassembly, for measuring an amount of torque transmitted through a torqueestablishing element in a multiple-ratio automatic transmission for anautomotive vehicle powertrain including a controller, a transmissionhousing, an input torque source, a torque output member, gearingdefining multiple torque flow paths from the input torque source to thetorque output member and a pressure actuated torque establishing elementfor establishing a gear configuration with a low speed ratio or a gearconfiguration with a higher speed ratio during a ratio upshift eventhaving a preparatory phase, a torque phase and an inertia phase,comprising: a core mounted on the transmission housing; and a loadsensor mounted on the core and positioned against a portion of thetorque establishing element whereby a portion of the amount of torquetransmitted through the torque establishing element travels through theload sensor and is measured by the load sensor assembly, wherein asignal representative of the amount of torque transmitted through thetorque establishing element is sent to the controller from the loadsensor.
 23. The load sensor assembly of claim 22, further comprising acover extending over the core and the load sensor.
 24. The load sensorassembly of claim 22, wherein the torque establishing element is a bandbrake including an anchor bracket and a band strap.